Hydraulic machine



March 23, 1954 E, K. BENEDEK HYDRAULIC MACHINE Filed Oct. 25, 1948 2, Sheets-Sheet l Oh wn INVENTOR.

BLEK K BENEDEK ,fi maw' Mal'fih 1954 E. K. BENEDEK 2,672,826

' HYDRAULIC MACHINE Filed Oct. 23, 3.948 '2 Sheets-Sheet 2 FIG 2 84 86 INVENTWZQ 6 YELEK K BENEQMQ FIG f 6 ra z /flmlwzw Patented Mar. 23, 1954 UNITED STATES @tlCE HYDRAULIQ MACHINE Application October 23, 1948, Serial No. 56,191

9 Claims. 1

This invention relates to rotary hydraulic ma chines and more particularly to hydraulic pumps or motors of the kind which include a rotary barrel formed with a plurality of circumferentially deployed radial cylinders in which reciprccacle pistons are mounted, and is a continuation-4m part of my copending application Serial Number 618,890, filed September 2'7, 194:5, now Patent No. 2,452,541, November 2, 1948. Hydraulic machines of this general class are well known in the art and have been used extensively under exacting conditions of fluid pressure and rotative speed. Although numerous improvements and refine ments have been contributed to the art in recent years, some dimculties remain to be eliminated or reduced in effect. One of the diniculties is that of maintaining proper lubrication of the parts which transmit radial thrust to the ciprocable pistons. Another undesirable characteristic of machines of this class heretofore pro vided has been their tendency to suck air into the cylinders resulting in noisy operation, and vibration frequently leading to break downs. In nearly all rotary hydraulic machines of the class referred to the driving torque is transmitted from i;

the cylinder barrel through the radial pistons themselves. At any stage of operation one-half of the pistons are operating on suction strokes, and in prior art constructions the suction pistons must be operated without the protection of pressure film of lubricant.

In the present instance the pistons are of the rigid T-head type which characterize the pump for their provision of a synchronized driving and driven connection between cylinder barrel and reactance structure. The arrangement is such that the piston proper is received and reciprocates in its respective cylinder while simultaneously the integral T-head, which is arranged at right angle to the axis of the piston and the reciprocation thereof, reciprocates in a tangential T-slot provided in the associated portion of the reactance ring, there being one T-slot for each piston crosshead.

The cylinder barrel or primary rotor and the reactance ring are disposed with a certain relative eccentricity e, and in operation each integral piston bodily carries out a planetating motion in the cylinder bore and the associated T-slot in such a manner that the piston movement, which will be the respective stroke of all the pistons, will compensate for the radial, while the motion of the T-head relative to and in its guiding T-slot, will compensate for the tangential component of the pump eccentricity. While each piston is reciprocating relative to its coacting primary and secondary rotors, it carries a prorated driving torque from the driving or primary, to the driven or secondary rotor. One difficulty experienced with pumps of the prior art is the lack of good and permanent alignment of the pistons and their respective crosshea-ds with their coacting primary and secondary rotors because the high speed and the heavy load impose a shock load upon the rotors via the pistons, the said shock load soon exhausts the capacity of the materials of the rotors, and loosens the T-slot carrying guide members of the reactance rotor. Then in a proportionate rate with the loosening up of the T-slot guides, the pistons themselves gradually wear egg shaped bores in their respective cylinders.

An object of the invention is to overcome these difficulties by providing an hydraulic machine of the class referred to in which novel rigid and reinforced reactance rings are formed in such a manner that an inner reactance ring is provided with a tangential series of straight sides, an outer polygon, concentrically surrounds the inner reactance ring and the polygonal sides of the inner ring are disposed in radial alignment with the respective polygonal sides of the outer reactance ring and the rings are spaced radially into an exact concentric position. By a simplified novel structure I provide not only permanent and indestructible spacing means radially between the polygonal rings, but by the same means, I clamp the rings into a reinforced permanent assembly between normally aligned parallel and axially separable side members. The clamping means are arranged parallel with the axis of rotation of the two main rotors, and are evenly distributed circumferentially of the reactance rotor, and between the two polygonal rings of the reinforced reactance ring. The axially separable side members are provided with radially aligned internal recesses, and concentric axial shoulders, which recesses will engage and support the inner and outer circular end shoulders respectively at both ends of the reactance polygons. Thus it will be seen, that all radial hydraulic piston load will be supported on the internal recessed shoulders of the axially separable side members. The function or my novel clamping and radial spacing means will be limited only to transmit axial clamping and preloading forces. In this respect, any radial shock load of the piston will hammer against the shoulders of the inner recesses and not on the coupling means. Thus by this distinct separation of the structure for radial shock load of the pistons, and the structure for spacing radially the polygonal faces of the inner and outer polygonal rings of the reactance structure, and for clamping the polygon rings respectively between recessed normally aligned faces of axially separable side members of the reactance ring, I succeeded to provide a durable, precision reactance structure, which has all the desired characteristics of a good reactance, for high speed heavy duty pumps of the above character.

A further object of the invention is to lock and secure rigidly the axially separable side members of the reactance against the polygonal rings of the reactance, because the axially separable side members are the supporting elements of the reactance for rotation, and the precision of holding and clamping the side members in assembly determines the center of rotation, as well as the functioning of the entire pump, not only during fixed stroke operations but during stroke control operations as well.

A further object or" the invention is to provide the spacing and novel clamping means of the reactance structure on a circle, substantially concentric with the axis or" the reactance, and by this simple geometrical relation, space the polygons radially by the clamping means, thus provided on a geometrical circle with their axis.

A further object of the invention is to provide a novel reactance structure, which will be able to absorb shock and vibration within the capacity of the pump pressures and eliminate breakage and fatigue phenomena within the pump.

A further object of this invention is to provide a novel reactance structure which includes two concentric radially spaced polygon rings, two axially separable side members, and axially preloaded spacing and clamping means, to assure a permanent, shock absorbing, reinforced reactance structure.

A further object of the invention is to provide an hydraulic machine including a pintle element and a surrounding cylinder barrel element and novel and improved means for sealing the clearance between the pintle and the cylinder barrel bore on opposite sides of registering pintle and cylinder barrel parts.

A further object of the invention is to provide an hydraulic machine including a pintle element and a rotary cylinder barrel element, and novel and improved means for supporting the cylinder barrel upon the pintle element with evenly distributed radial clearance, and by placing said novel and improved means so close to the pintle and barrel parts as to provide only sufficient viscous seal between pintle and barrel, and minimum supporting span for the pintle and barrel. A further object is to provide improved bearing means between pintle and barrel whereby the radial clearance between the viscous seal surfaces of the elements will be evenly distributed and thereby the pressure slip will be minimized, and

simultaneously the novel bearing elements will be disposed so close to the main inner pump parts as to minimize elastic deflection of the pintle and the barrel, and thereby make the operation of my novel bearing means possible.

Other objects will become apparent from a reading of the following description, the appended claims, and the accompanying drawings,

in which:

Figure l is a vertical, longitudinal main section of the pump, showing one embodiment of the present invention.

Figure 2 is a transverse section, taken on line 22 of Figure 1, showing partly in section, partly in elevation, the radial series of piston and cylinder assemblies, constructed in accordance with this invention.

Figure 3 is a section taken on line 33 of Figure 1.

Figure 4 is a fragmentary enlargement of the T-crosshead receiving portion of my novel reactance means.

Figure 5 is an elevation of my clamping and spacing means.

Figure 6 is a fragmentary enlargement of the clamping and spacing receiving portion of my novel reactance means.

Figure 7 is a section taken through line 1-! in Figure 2, showing one polygon of my novel eccentric ring, with its associated piston and piston receiving aperture.

Figures 1 to '7 illustrate the invention as being embodied in a rotary radial cylinder and piston pump of the kind employing crossheads or T- heads for thrust transmitting connections instead of the rolling pins described with reference to the constructions shown in Figures 1 to 9, inclusive, of the parent application, Serial Number 618,890, now Patent No. 2,452,5dl, November 2, 1948.

The pump shown in Figures 1 to 7 include a stationary casing F including plates 59 secured to the ends of the casing proper by means of screws 60. A pintle G is formed with a valve portion El and with a mounting portion 62 of larger diameter which is fixedly mounted in a boss 83 on the left hand end plate 59. The pintle is formed with passages, the outer ends of which are indicated at 64 and 65. The passages 64E5 extend from the indicated outer ends longitudinally through the pintle and communicate respectively with pintle ports 68 and ti.

Mounted for rotation about the pintle G is a cylinder barrel H formed with a plurality of circumferentially deployed radial cylinders 68 adapted to communicate with the pintle ports 66 and 61 by means of cylinder ports 69. Preferably the rotative mounting of the cylinder barrel is provided by means of needle bearings H interposed between cylinder barrel H and the pintle portion 6|. In order to supplement the radial load mounting of the barrel and to hold the barrel against axial shifting, a cap 12 on the end of a shaft it is secured to the cylinder barrel by screws I4 and is journaled in the right hand casing end plate 59 by means of bearings'15 held in place by an end cap '16 secured to the right hand end plate 59. The inner races of the bearing 75 are clamped against the shaft cap M by a nut ll. Pistons it are mounted for reciprocation in the cylinders 58.

Referring now more particularly to the construction of the reactance means, it will be seen, that the reactance structure includes the two concentric polygons it and St as in Figure 1. Since there must be an operating clearance between the parallel coasting polygonal faces of the rings l9 and it is obvious that two important things are necessary to obtain such operating clearance. First, each pair of coasting polygon faces must be absolutely parallel between themselves, and second, the radial and parallel distance between coordinated faces must be exactly the same between all paired faces. This second requirement, as it will be seen hereinafter, is the same as the requirement of concentricity of the inner l9 and the outer 89 polygons. According to the spirit of the invention, I solved these two simultaneous requirements by a simple constructive method. Namely, on the inside radial faces of side members tl--8| I provided normally aligned concentric recesses, see Figure 4, in such a manner that each recess determines an inner and an outer recess shoul der. The polygons l9 and 80 are provided with mating circular shoulders to mate the inner 2t and outer lit circular recess shoulder of the side members 3 lti. The arrangement is such, see Figure 6, that the total radial width of the two polygons is and 8t) plus the diameter of coupling bolts 82 is equal or slightly greater than the radial dimension of the recess shoulder til and 85 as in Figure 4. In this manner when coupling bolts, as in Figure 5, will be inserted and drawn tight between the rings is and at on the one hand and between the side members 8l-ili on the other, the bolts 82 will radially separate the rings is and so and press them against the mating recess shoulder ill and Bi of the side members Bl-8l radially, and the two parallel circular or annular ends of the spaced polygons will be axially forced into and seated in tight abutment in the recess of the side members El-fil as shown in Figure 4. Thus, while in Figure 4 the polygons are radially separated and pressed radially against the shoulders ti and 8& of the load transmitting recesses, piston crosshead it is perfectly free radially and has a predetermined radial operating clearance of a predetermined amount, consistent with the size of the crosshead, the speed and the pressure of the pump. Also, it will be seen that when the above two requirements for alignment of the polygons l8 and 3d are fulfilled, the apex points or lines of both polygons will lie on concentric parallel circles and each pair of radially aligned inner and outer apex points will lie on a com-- mon line, which goes through the center of rota tion of the polygons, which now in assembly corrstitute the reactance assembly. Thus the positive and radial spacing of the polygons l9 and ell by means of coupling bolts M as shown in Figure 6, provides the permanent and free running clearance for the thrust transmitting T-heads it, between radially aligned polygonal faces of the resultant reactance, and thus provides free reciprocation for the pistons in both radial and tangential directions. While the free reciprocation of the piston is radially assures good pumping action, the free reciprocation of the integral Y-head crosshead l8 tangentially means efiicient torque transmission between cylinder barrel H and reactance assembly ll -.'--8 i The efficient torque transmission further means,

torque alone from the cylinder barrel to the reactance ring.

The structural as well as the manufacturing advantages of the concentric reactance rings l9 and till cannot be emphasized enough. Each ring has, individually, to transmit tremendous radial and torque (tangential) forces and at the same time remain undistorted, fixed, in its assembly, not only relative to itself, but relative to the other, inner to the outer, outer to the inner reactance ring. Additionally, the reactance rings "ill and must remain undistorted with respect to the side members 8i3l also. No known form or structural shape is so rigid, so compact and so fit to fulfill the above complex requirements as a tubular or cylindrical memher or members such as it or tel. When it comes to the manufacturing of such members as it or 80 out of seamless steel tubing or ring shaped castings, by cutting them to equal length axially and broaching or milling the flats on the inside of member 89] and on the outside of there is no milling or broaching operation which could be expedited with greater accuracy and at less expense than these operations forming the polygonal reaction or thrust faces on the reactance tubular ring iii or 89. Furthermore, once these faces are milled, broached or finished on the rings, they remain there in proper, undisturbable relative positions with respect to each other and other associated working parts. Because the rings are made out of one piece of material, they are integral and rigid in themselves, and stronger than any composite, inulti-part structure of the prior'art.

It must be emphasized that the location of coupling bolts 82 in the side faces of side members tliii is a simple and inexpensive mechanical operation, as these holes will be bored according to a plate pattern which will located and fixed into the inside recesses of the side members til-ti, and with respect to the recess shoulders ti and 8t respectively which will finally control the location of the bolts 82.

While the positive and uniform radial running clearance between the T-heads i3 and the coasting polygonal faces of rings '1?) and it gives better lubrication to the heavy torque transmitting reciprocation of these parts, Figure 4 shows that a circular recess as at til is provided all around the fluid tight outer polygon til for lubrication. This liquid will be there and ready for lubricating, irrespective of the provisions of the parent case of this application, Serial No. 618,890, filed September 27, 1945, now Patent No. 2,452,541, November 2, 1948.

Another great advantage of utilizing the above coupling means 62, see Figure 5, lies in the possibility of preloading, that is, in pulling very tight the side faces of side members 5l--Bi against the end faces of the polygons l9 and 8t and possibly tighter, than the stresses which will be caused in these bolts by the bending action oi the hydraulic shock load of the several pistons. The method of preloading the coupling or clamping bolts 82 occurs in various manners. One preferred method comprises the heating of the bolts prior to the assembly of the reactance with the piston and cylinder subassemblies, to a couple hundreds of degrees Fahrenheit above the shop temperature. The substantially long clamping bolts, when heated to 200 F. excess temperature will expand sufficiently in length so that in assembly the heat expanded bolt will shrink and be pulled tight against side members lib-Si and the rigid concentric rings 9 and til After the entire assembly will cool down to shop temperature, the bolts contract and cause a preload force in the bolts longitudinally, in addition to the tension pull of the assembly. One feature of this assembly is that both the side members and the polygons are adaptable for such preload.

The pressure loaded pistons on the one side of the reactance structure cause a bending moment which imparts compression forces to the bolts 82, while on the opposite side of the reactance this same bending moment imparts tension forces to the bolts of its half circle side. lhese alternating stresses, as the reactance rotor rotates under very high speed, change during each revolution at least once, from tension to compression, it will be seen that a good preload of the bolts 82 not only will produce a rigid tight structure, but eventually will minimize the breakage of the bolts 82, or prevent their getting loose under the great frequency of the alternating shock loads.

A further outstanding contribution of the novel clamping means 82 to the construction of this invention lies in its ability to carry heavy duty, high frequency shock load. Its simplicity and manufacturing advantages will be appreciated only when one compares this reactance structure with the solid, forged, heavy one piece reactance of the prior art, which almost was prohibitive to make. It will be seen that while rings l9 and Bi? may be made equally well out of forged rings or steel tubing, which are both available in large quantities and at commercial prices, an integral forging would cost several times as much to forge, machine and assemble, and yet many advantages of the present structime would be missing. The disposition of the reinforcing bars 82 on a large diameter bolt cirole adds greatly to the stifiness of the Stl'ZC'GUl'E, and very little to the Weight of it. Yet the axially separable side members ill-ti will permit precision machining and lapping of the polygonal faces, while similar faces in an. unseparable member are almost inaccessible to any production operation.

The new features and advantages rendered by the coupling means 82 to this combination are one or the basic distinctions between this application and the structure of my prior Patent 2,373,449, April 10, 1945, which is lacking in these new features and advantages.

The thrust transmitting connections for the pistons it are in the form of T-heads or crossheads 73 extending tangentially with respect to the pintle axis. The inner faces of the heads TB have sliding contact with tangential faces on an inner reactance polygon l9; and the outer faces of the heads it slidably engage tangential faces on an outer reactance polygon til. The two polygons l9 and til are positioned between reactance rings 8i8i drawn into clamping engagement with the polygons "is and as by means of bolts 82. The reactance assembly comprising the polygons is and 80 and the rings 2i is mounted for rotation by bearings 83 held in an outer reactance member 85 by means of snap rings 85. The outer reactance member t l and with it the inner reactance assembly are mounted to shift transversely of the pintle axis by means of pads 85 on the outer reactance ring 8 3 and fixed pads 8? on the casing F. The reactance structure may be shifted as a unit by any suitable mechanism connected to the ring as by means of shifting rods 83.

Considering now the outside or reactance rotor as a self contained assembly, it will be understood that this rotor structure as a subassembly includes side members ill-8i, inner and outer concentric polygons 'l8i! and the plurality of clamping bolts 82. In considering the specific merits of each of the elements of this sub-combination, it will be seen, that the side members 8l--8l are simple, rigid, and possess a geometry which enables them to be parallel, to have one common axis of rotation, and to provide the normally aligned opposed recesses of the polygons 1S--S0 in exact concentricity with the axial flanges and the center of rotation of the members 8l-8I in one solid non-yielding body. As it will be seen hereinafter, for the purposes of this invention and for the proper functioning of the polygons l9 and 83, as well as for the proper alignment of the T-heads T8 in between the polygons l9 and 89, such rigid unyielding structure and workable geometry is an absolute pre-requisite. This refers to both the outside and inside lips of the side members iii-8|, which lips confine, engage and axially align the polygons i9 and and with them the working polygonal thrust faces for piston crossheads l8l8- 18, etc. The axial alignment of the groove-ways of the side members til-3| makes the radial alignment not only for the polygons is and 8% possible, but also makes possible the radial alignment of the end flanges of the side members 8l- 8l with their respective bearings 33-83. All these excellent possibilities and potentialities however for the provision of a rigid, simplified rotor structure would remain theoretical, until after the provision for a similarly excellent means and method for securing the parts into a safe and permanent working assembly, which assembly has to withstand many abuses such as heavy alternating and shock loads, vibrations and even eventual skewing and wobbling, if the assembly would get loose. Such adequate means and provision is found in the application of a plurality of my novel rigid, simple and inexpensive clamping and reinforcing means 82. These clamping means 82 provide the following three fold basic improvements to my reactance rotor subcombination. First, they clamp the rotor side members against the radially aligned and abutting ends of the polygons it and 8B. In this connection the polygons serve as rigid spacers axially from end to end and between the radial faces of the respective grooves of the side members. Second, the coupling means i222 operate for the polygons 198il as spacers radially, in that they separate the polygons radially, because they are disposed between the polygons in such a mannor that they are in direct spacing contact with the confronting surfaces of the polygons. But because the bolts 532 are further disposed in the confronting corners of the polygons they are positively retaining the polygons against transverse rotation so that each pair of coasting polygonal faces remain parallel to each other in addition to spacing and fixedly securing them radially with respect to each other and to the side members respectively. All these combinations were possible because of the simplified geometry of the side members 81-3! and the polygons iii-did respectively, and the fact that the screw holes in the side members 8i-8i for the clamping means 82 could be provided with sufficient accuracy in conventional machine tool operations. See also Fig. l and Fig. 2. Thirdly, the proper geometry of the subcombination again makes the fastening and clamping operation of the bolts 82 such that the subassembly when properly tightened and locked by some conventional method, will be so reinforced, strong, rigid and secured, that load and torque transmission of the several pistons iii- J8 cannot dis.- turb either the assembly or its proper functioning. Incidentally the coupling or clamping bolts 82 save tremendous weight through their strength giving arrangements.

From an hydraulic viewpoint there is a very important relationship between the hydrostatic pressure slip and the amount of the radial clearance between closely fitted and lapped pintle and barrel surfaces. When this lapped clearance is of the magnitude of one half to one thousandth of an inch per one inch pintle diameter, it is often called a capillary clearance, or the sealing effect itself a viscous seal. In capillary seals between pintle and barrel it is very important that the clearance be evenly distributed around the entire periphery of the pintle and the barrel bore respectively, in order to obtain the maximum seal or the minimum amount of slip. Pressure slip is a definite hydraulic loss which reduces the overall efficiency of the pump.

While several of my U. S. patents, such as for instance 2,181,732, proposes and uses needle roller bearings between pintle and barrel, none of them teaches the importance of the nature of the capillary clearance around the cylinder and pintle ports 55 and Eli respectively. In 2,101,732, fur thermore, needle roller bearings are employed only as pilot bearings for the pintle, since the barrel is fully and solely supported independently of the pintle. The present application, however, employs the needle roller bearings ll--ll as main bearings between pintle and barrel, and the barrel is supported actually on said needle roller bearings, which rest against the rigid pintle. As for the amount of the slip, the item portant matter is the distribution of the radial clearance between pintle and cylinder barrel quite at the zone of the ports 85 and 69 not so much thereafter toward the ends of the pintle l.

I, therefore, in the present application propose not only preloaded centering means such as preloaded needle roller bearings at both sides for the piston and cylinder assemblies, but I apply each bearing structure as close to the pintle ports as possible, to be sure that elastic deflection and offset clearance is minimized to i a practical minimum. Thus by maintaining concentric and evenly distributed clearance I reduce slip, reduce capillary seal length, and increase operating efficiency of the pump. One main reason for obtaining reduced slip lies in the viscosity of the working fluid. As long as the clearance is small and capillary and evenly distributed all around the pintle, the viscosity of the oil resists flow fluid, to a great extent, and the slip is a minimum. eccentric, however, and the clearance is zero on one side and 2C on the other side (if C is the radial clearance when pintle and barrel are concentric), '-aiscosity is losing sealing effect on the large clearance side and the slip becomes excessive and amounts to two to three times as much as the slip at the same pressure but with the clearance uniform, that is, C all around.

To comply with this inventive principle, I counterbore the ends of the barrel and interpose capillary, full complement needle rolle assemblies as at li 'll between pintle and barrel, see also Figure 3, in such a manner that irrespective of the distribution of the working pressure fluid in peripheral clearance C, between pintle G and the bore of the barrel the clearance C will remain unaffected and. unchanged by the maximum unbalanced load, say 3060 p. s. 1. Thus, outside or the close disposition of the bearings H ll to the cylinder ports til, I assemble the When pintle and barrel become Iii needle roller assemblies ll-ll with suflicient radial preload to keep th operating or capillary clearance C at a predetermined constant uniform value. The actual valv of C is about onehalf to one thousandth of an inch per one inch diameter of th pintle. Thus a two inch diameter pintle needs from one to two thousandth of an inch radial clearance, between closely finished concentric members. Since the slip is increasing very rapidly with C, it is obvious that the cutting of G into half means exactly the cutting of the length of the seal into one-half, that is the distance or the bearings from the cylinder port 69 will be one-half as short as before. The great advantage of cutting the length of the pintle and barrel seal can be appreciated only when the amount of the deflection of the unsupported span of the pintle is wholly considered. In my U. S. Patent 2,166,717, I already proposed a minimum pintle seal length by the viscosity of the working fluid, but I did not propose there the use of the centering means for the purpose of maintaining a constant uniform clearance C in combination with a short seal. Thus in such a construction, when the pintle bends, and the clearance become upset, the slip may become unnecessarily high, and two to three times of its normal value.

As shown in Figure 1 and Figure 3, elongated rollers 'li'll preferably with their tool-steel race are interposed between pintle and barrel end flanges under predetermined radial pressure, to preload the bearing structure for the maximum bearing load, and thus eliminate any further load deformation between rollers "ll, pintle G, and coacting bearing races. Such a combination will be set to run substantially at a constant clearance C under all load and speed conditions. The method of preloading needle roller assemblies ll--"ll may vary according to the use and purpose of such bearings, In the present embodiment I prefer a simple and inexpensive method for this purpose. Actually I make the out" side diameters of the barrel counterbores at both ends of the barrel smaller than the mating diameters of the bearings or bearing assemblies lli l, in such a manner that in assembly when the bearings ll--ll will be assembled in their respective ccunterbores, there will be a definite interference fit between the diameter of the oouriterbore and the diameter of the respective hearing or bearing assemblies. Next, I mak the corresponding portions 6 l-lil of pintle G greater in diameter than the inner diameters of the respective bearings or bearing assemblies 'Il7l. In this manner when the hearings will b assembled upon the pintle portions 6l6l there will be a positive interference fit between the ouside diameter of the pintle portion iii and the inside diameters of bearings or hearing assemblies IL- ll The interference fit, predetermined, will give the required preload of the bearings. In practice the assembly of the barrel II and the bearings l l-ll will occur by the heating of the barrel II, or by the cooling of the pintle G to a predetermined degree. After the barrel is heated, the radially oversized bearings ll-H, in spite of the interference fit between the barrel bore and the bearings themselves, will readily slip into their respective places. After the preloaded parts will cool down to shop temperature, the bearings iii-ll will be preloaded radially between the pintles 6-6? and barrel bores I iil respectively.

Needle rollers of the bearings 'H-ll arefree 1 1 otherwise axially and circumferentially, as there is a total circumferential clearance of about one roller diameter for the entire bearing, and a total of ten thousandths of an inch (.OlG) end clearance between race lips and needle ends. In such a preloaded needle roller structure the needles will roll freely and thus render an antifriction bearing means with practically no runout for supporting the pintle relative to the barrel. It will be seen that the rigid pintle and the right hand supported barrel upon radial and thrust bearings 15 will coact in a unique manner. Bearing 15 will support the right end of the barrel and the free end of the pintle, while the walledin pintle supports the left hand of the barrel and the left end of its pintle portion M. An important advantage of the cageless capillary needle rollers lies in their ability to carry large radial load such as ten to twenty thousand pounds per projected square inch area, while other bearings, even such heavy duty types as the tapered roller bearing 15, can carry only one third of the above specific load. Additionally, the clastic deformation and radial runout of the small elongated rollers is so small, as compared to the large bearings of 75 that they constitute ideal centering means for a uniform and capillary clearance 0 between pintle and barrel, and thus cooperate to seal the working fluid by its viscosity and coacting uniform capillary clearance. Clearance C is naturally small and capillary, but its actual amount used in this connection was given as in the neighborhood of one to two thousandths of an inch, while the actual length of the capillary seal at each side of the piston and cylinder ports 69 depends entirely on the amount of C as hereinabove outlined, also depends on the ability of centering means i i-'ii to maintain C at a constant predetermined value.

When precision workmanship reduces C to its lower value, that is to one-half of one thousandths per inch pintle diameter, and the bearings ll-H may sustain the constant clearance at one-half of one thousandths of an inch, the

seal length at each side of port 89 may range between one fourth to one-half of the pintle diameter. Improved finish and material may further reduce this minimum seal, and the bearings II-ll may be spaced even at a lesser distance from port 69.

In order to introduce lubricating or sealing fluid into the chamber I, fluid which works along the pintle in the clearance between the pintle and the cylinder barrel bore is collected in a chamber at 92 at the free end of the pintle and is led past the adjacent bearings H and through one or more passages 93 extending through the cylinder barrel and opening into the chamber 1. In operation, th slip fluid passing through the clearance between the pintle and the cylinder barrel bore will be conducted to the chamber I, thereby filling the chamber and maintaining it full so as to lubricate the thrust-transmitting parts effectively and also to seal the outer ends of the cylinders against, the ingress of air.

The importance of the above advantages can be appreciated only when it is well understood that the operation of a radial pump of this character is accompanied by periodic shock loads and vibrations. further amplification of these vibrations with the resulting effects of fatigue, breakage and loose part connections. The above novel structure, however, prevents a priori such destructive vibrations, and the elastic preload minimizes the The, finite number of pistons causes possibility that those vibrations come into synchronism and produce critical phenomena which are apt to cause substantial danger to the safety of operation of machines of this character.

The constructions disclosed embody the invention in the forms now preferred, but it will be understood that changes maybe made without departing from the invention as defined in the claims.

I claim:

1. A pump or motor comprising a casing, primary and secondary rotors within said casing, a radial series of piston and cylinder assemblies interposed between said rotors, said secondary rotor comprising complementary recessed parallel side members, a pair of concentric polygons arranged between said side members, clamping means disposed between said polygons, said polygons being clamped end to end against said side members and said clamping means respectively, whereby the assembly of said polygons, side members and clamping means constitute a reinforced reactance structure, said clamping means being coextensive axially with the axial length of said reinforced structure, and crosshead means on said assemblies engaging said polygons.

2. In a pump or motor, primary and secondary rotors mounted for concurrent rotation about parallel axes and movable relatively during rotation, piston and cylinder assemblies including crossheads arranged between the rotors, a pair of concentric reactance rings having tangentially parallel and radially opposed plain bearing surfaces engaging said assemblies, and means for securing said reactance rings to said secondary rotor, said securing means extending between the opposed bearing surfaces of said reactance rings and contacting the opposed inside and outside polygon of said reactance rings, whereby the polygons of said rings constitute a tangential series of crosshead receiving guideways.

In a pump or motor, primary and secondary rotors mounted for concurrent rotation about parallel axes and movable relatively during rotation; piston and cylinder assemblies including crossheads arranged between the rotors, a pair of concentric polygons engaging said assemblies, and means securing said polygons to said secondary rotor, said securing means extending between the sides of said concentric polygons and engaging the radially aligned sides of the polygons, whereby the assembly of said polygons and said securing means is forming polygon shaped tangential groove-ways, the radial distance be tween the opposed faces of each groove-way being greater than the thickness of the associated crosshead to thereby allow free reciprocation of the crossheads in said groove way;

4. In a pump or motor, primary and secondary rotors rotatable about individual axes, piston and cylinder assemblies connecting the rotors and including crcssheads, said secondary rotor including side members, a pair of concentric polygons interposed between said side members, means for ciamping the side members against the ends of polygons and against each other, said means comprising clamp members positioned one between each adjacent side of said polygons, and

polygons directly engaging said clamp members and being positively held against displacement by such engagement.

5. In a rotary radial piston pump or motor, a rotatable barrelhaving a plurality of radial cylinders, valve means for the cylinders, a rotary reactance in eccentric relation to the barrel, and

surrounding the same; crossneads on the pistons, operatively connected to said rotary reactance, said reactance including a set of paired polygonal rings, each opposed pair of polygonal faces and the intervening crosshead constituting a thrust transmitting reactance bearing assembly for said piston crosshead, the radial spacing of said faces being greater than the radial dimension of said crosshead, and means other than the crosshead engaging said faces to maintain the radial spacing and radial alignment of said faces.

6. In a pump or motor, primary and secondary rotors mounted for concurrent rotation about parallel axes and movable relatively during rotation, piston and cylinder assemblies including T- shaped crossheads arranged between the rotors, a pair of concentric reactance rings having parallel and radially opposed plain bearing surfaces engaging the T-heads of said assemblies, a pair of planetating lubricant-retaining rings operatively interposed between the rotors and said pair of reactance rings for retaining lubricating fluid for said polygonal faces and coacting T- heads between the rotors, and means other than the T-heads to space said polygons concentrically to each other and with their opposed plain bearing surfaces in parallel relation, thereby maintaining said polygons in proper concentric and angular relation during the operation of the pump or motor.

7. In a pump or motor, primary and secondary rotors rotatable about individual axes and movable relatively during rotation, radial piston and cylinder assemblies connecting the rotors including crossheads, said secondary rotor comprising parallel side members having opposed circular grooves axially aligned and concentric with the axis of rotation of the secondary rotor, a reactance ring clamped between the rotor side members and having inwardly facing polygonal reaction faces to slidably support the crossheads and having circular reaction shoulders at both ends engaging the grooves at the rotor side members, and clamping means for tightly holding said reactance ring in compressed position between said side members, said clamping means comprising clamping bolts positioned equidistantly about the secondary rotor axis, one between each pair of adjacent polygonal faces and being coextensive axially with said reactance ring, said polygonal cular grooves axially aligned with each other, a pair of concentric reactance rings clamped between the rotor side members and having radially aligned opposed parallel, polygonally arranged, reaction faces to slidably receive the crossheads, each of said pair of concentric reactance rings having circular shoulders at both ends thereof engaging the grooves of the side members, clamping means for clamping the pair of concentric reactance rings firmly between said side members, said clamping means comprising clamping bolts positioned about the secondary rotor axis between each adjacent pair of polygonal faces and passing axially completely through said concentric reactance rings, and means on each of said pair of concentric reactance rings directly engaging said bolts for positively holding said concentric reactance rings against displacement relative to said side members.

9. In a pump or motor including a cylinder barrel, a cylindrical valve pintle member, a cylindrical barrel member mounted on said valve pintle member for rotation relative thereto, radially disposed cylinders in said cylindrical barrel member arranged to communicate with said valve pintle member, pistons arranged to reciprocate in said cylinders, means for reciprocating said pistons and radially preloaded elongated needle roller bearings mounted between said members and means associated with at least one of said members for permitting movement of that member axially relative to said needle bearings whereby said members may move axially relative to each other to compensate for deflection of said pintle resulting from the loads applied by the reciprocation of said pistons.

ELEK K. BENEDEK.

References Cited in the file of this patent UNITED STATES PATENTS Number Name Date 1,308,436 Maw July 1, 1919 2,000,271 Benedek May 7, 1935 2,024,420 Benedek Dec. 17, 1935 2,041,172 Ernst May 19, 1936 2,130,298 Ernst Sept. 13, 1938 2,236,666 Benedek Apr. 1, 1941 2,371,078 Summers Mar. 6, 1945 2,373,449 Benedek Apr. 10, 1945 2,374,592 Ernst Apr. 24, 1945 2,426,588 Benedek Sept. 2, 1947 OTHER REFERENCES Machine Tool Applications Publication issued by Norma-Hoffman Bearing Corp., received April 20, 1937; page 5 only. (Copy in Division 9, 103-161.) 

